RAILWAY VEHICLE VIBRATION DAMPING DEVICE

Abstract

A railway vehicle vibration damping device includes an actuator interposed between a vehicle body and a truck and capable of exerting a control force; and a controller configured to obtain a target control force for reducing a vibration of the vehicle body based on a lateral acceleration of the vehicle body, and the controller includes a bandpass filter configured to extract a low frequency control force, which is a frequency component lower than a resonant frequency of the vehicle body, and a correcting part configured to correct a target control force based on the low frequency control force.

Description

TECHNICAL FIELD

The present invention relates to an improved railway vehicle vibration damping device.

BACKGROUND ART

A railway vehicle includes a railway vehicle vibration damping device including a double-acting actuator interposed between a vehicle body and a truck, and a controller configured to control the actuator and is configured to reduce a vibration in a left-and-right direction with respect to a direction of travel of the vehicle body. The railway vehicle vibration damping device inputs an acceleration in a left-and-right direction of the vehicle body detected by an acceleration sensor to a controller, controls an actuator based on an acceleration feedback, and thus is capable of controlling a left-and-right movement of the vehicle body.

When the railway vehicle travels a curve zone, a centrifugal acceleration referred to as a stationary acceleration acts on the vehicle body. Therefore, the acceleration detected by the acceleration sensor includes the stationary acceleration in addition to an acceleration entering from the truck side and vibrating the vehicle body in the lateral direction.

A main reason that deteriorates riding comfort is the vibration of the vehicle body in the lateral direction, and thus if an attempt is made to make the actuator exert a force resisting the stationary acceleration, the actuator is obliged to exert an excessive force. Consequently, an excessive load is applied to a motor that drives a pump as a drive source for the actuator. Accordingly, as disclosed in JP2012-245926A, the railway vehicle vibration damping device is configured to perform a process of removing the stationary acceleration from an acceleration detected by an acceleration sensor, control the actuator based on an acceleration feedback, and then reduce the vibration to improve riding comfort.

SUMMARY OF THE INVENTION

A frequency band of the stationary acceleration is on the order of 0.3 Hz, and a resonance frequency band of the vehicle body is generally on the order of 1 Hz to 1.5 Hz, and both frequency bands are very close to each other. Removal of the stationary acceleration is insufficient when a lower cutoff frequency of a bandpass filter configured to remove the stationary acceleration is too low, while effects of vibration damping is reduced when the lower cutoff frequency of a band-pass filter is too high because vibration components to be reduced are also removed. Accordingly, the cutoff frequency of the bandpass filter is obliged to be set in a very limited range between the frequency band of the stationary acceleration and the resonance frequency band of the vehicle body.

When the railway vehicle travels from a straight zone to a relaxation curve zone provided between the straight zone and the stationary curve zone, a centrifugal acceleration acts on the vehicle body. The frequency band of vibration caused by the centrifugal acceleration is higher than the frequency band of the stationary acceleration during traveling in the stationary curve zone, but is lower than the resonance frequency band of the vehicle body.

With the setting of the cutoff frequency of the filter as described above, the centrifugal acceleration during the travel in the relaxation curve zone cannot be removed. Therefore, the railway vehicle vibration damping device in the related art performs control for making the actuator generate a control force for reducing the cutoff frequency.

As described above, since a force for pushing the vehicle body against the centrifugal acceleration is significantly large, the actuator needs to exert a large force, and in railway vehicle vibration damping device of the related art, a limiter that avoids an over loading of the motor is provided and thus controls the controlling force. When the control force is limited in this manner, the vibration of the vehicle body entering from the truck, which is essentially to be reduced, cannot be reduced, and consequently, riding comfort of the vehicle is degraded.

An upper limit of the control force of the actuator may be increased simply by upsizing the motor. However, it may result in a cost increase of the railway vehicle vibration damping device, and mountability on the railway vehicle may be impaired.

It is an object of the present invention to provide a railway vehicle vibration damping device capable of improving riding comfort during the travel in the relaxation curve zone without impairing cost and mountability on the railway vehicle.

A railway vehicle vibration damping device of the present invention includes: an actuator interposed between a vehicle body and a truck of a railway vehicle and capable of exerting a control force; and a controller configured to control the actuator by obtaining a target control force for reducing a vibration of the vehicle body based on lateral acceleration of the vehicle body, wherein the controller includes a filter configured to extract a low frequency control force, the low frequency control force being a frequency component lower than a resonant frequency of the vehicle body in the target control force and a correcting part configured to correct the target control force based on the low frequency control force.

BRIEF DESCRIPTION OF DRAWINGS

FIG. 1 illustrates a cross-sectional view of a railway vehicle provided with a railway vehicle vibration damping device according to an embodiment.

FIG. 2 illustrates a detailed drawing of an actuator.

FIG. 3 is a control block diagram of a controller in the railway vehicle vibration damping device according to the embodiment.

FIG. 4 is a control block diagram of a correction part in the railway vehicle vibration damping device according to a first embodiment.

FIG. 5 illustrates a drawing explaining a change in correction gain.

FIG. 6 illustrates a flowchart illustrating a process procedure in the controller of the first embodiment.

DESCRIPTION OF EMBODIMENTS

The present invention will be described below based on embodiments illustrated in the drawings. A railway vehicle vibration damping device 1 according to an embodiment is used as a vibration damping device for a vehicle body B of a railway vehicle, and includes actuators A to be interposed between the vehicle body B and a truck T as a pair, and a controller C configured to control the actuators as illustrated in FIG. 1.

Particularly, the actuators A, more specifically, in a case of a railway vehicle, are coupled to pins P suspended downward from the vehicle body B and are interposed between the vehicle body B and the truck T side-by-side in pair. The truck T retains wheels W in such a manner as to be freely rotatable, and springs S, S, are interposed between the vehicle body B and the truck T, and the vehicle body B is resiliently supported to allow the vehicle body B to move in a lateral direction with respect to the truck T.

The actuators A are basically configured to be actively controlled and reduce lateral horizontal vibration of the vehicle body B in a lateral horizontal direction with respect to a direction of travel of the vehicle. The controller C is configured to control the actuator A and reduce a vibration of the vehicle body B in the lateral direction.

The controller C in this example detects a lateral acceleration α of the vehicle body B in the lateral horizontal direction with respect to the direction of travel of the vehicle during control for reducing vibration of the vehicle body B. The controller C then obtains a target control force Fref that the actuators A are to be generated based on the lateral acceleration α, causes the respective actuators A to generate a thrust force according to the target control force Fref and reduces the lateral vibration of the vehicle body B.

Subsequently, detailed configuration of the actuators A will be described. The actuators A have the same configuration. In the illustration, two each of the actuators A are provided per truck T. However, a configuration of the truck T having only one actuator A is also applicable. Alternatively, one each of the controller C may be provided for each of the actuators A.

As illustrated in FIG. 2, the actuator A in this example includes a cylinder body Cy including a cylinder 2 coupled to one of the vehicle body B and the truck T of the railway vehicle, a piston 3 to be slidably inserted into the cylinder 2, a rod 4 inserted into the cylinder 2 and coupled to the piston 3 and other one of the vehicle body B and the truck T and being configured to be extendable and contractable, a rod-side chamber 5 and a piston-side chamber 6 defined in the cylinder 2 and separated by the piston 3; and in addition, a tank 7 storing a hydraulic oil; a pump 12 capable of sucking the hydraulic oil from the tank 7 and supplying the hydraulic oil to the rod-side chamber 5; a motor 15 configured to drive the pump 12; and a liquid pressure circuit HC configured to control switching between the extension and the contraction and a thrust force of the cylinder body Cy. The actuator A is configured as a single rod actuator.

The hydraulic oil as a hydraulic liquid in this example is filled in the rod-side chamber 5 and the piston-side chamber 6, and a gas is also filled in the tank 7 in addition to the hydraulic oil. The tank 7 is not necessary to be brought into a compressed state by being filled with a compressed gas. The hydraulic liquid may be other liquid other than the hydraulic oil.

The liquid pressure circuit HC includes a first opening-and-closing valve 9 provided in the middle of a first passage 8 communicating the rod-side chamber 5 and the piston-side chamber 6, and a second opening-and-closing valve 11 provided in the middle of a second passage 10 communicating the piston-side chamber 6 and the tank 7.

Basically, the cylinder body Cy is extended when the pump 12 is driven with the first passage 8 in a communicating state by the first opening-and-closing valve 9 and the second opening-and-closing valve 11 closed, and the cylinder body Cy contracts when the pump 12 is driven in a state in which the second passage 10 is brought into a communicating state by the second opening-and-closing valve 11 and the first opening-and-closing valve 9 is closed.

Respective parts of the actuator A will be described in detail. The cylinder 2 has a cylindrical shape, and a right end of the cylinder 2 in FIG. 2 is closed by a lid 13, and a left end in FIG. 2 is provided with a rod guide 14 having an annular shape. The rod 4 is slidably inserted into the rod guide 14, and the rod 4 is slidably inserted into the cylinder 2. One end of the rod 4 projects out of the cylinder 2, and the other end of the rod 4 in the cylinder 2 is coupled to the piston 3. The piston 3 is slidably inserted into the cylinder 2.

Note that a space between an outer periphery of the rod guide 14 and the cylinder 2 is sealed with a sealing member, not illustrated, and thus the interior of the cylinder 2 is maintained in a sealed state. The hydraulic oil is filled in the rod-side chamber 5 and the piston-side chamber 6 in defined in the cylinder 2 by being separated by the piston 3 as described above.

In the case of the cylinder body Cy, a cross-sectional area of the rod 4 is half a cross-sectional area of the piston 3, and a pressure receiving area on the rod-side chamber 5 side of the piston 3 is half a pressure receiving area on the piston-side chamber 6. Accordingly, when the pressures in the rod-side chamber 5 during an extending operation and a contracting operation are equalized, thrust forces generated both during extension and contraction are equalized and the amount of hydraulic oil with respect to the amount of displacement of the cylinder body Cy both during extension and contraction are equalized correspondingly.

More specifically, for the extending operation of the cylinder body Cy, the rod-side chamber 5 and the piston-side chamber 6 are kept in the communicating state. Consequently, the pressures in the rod-side chamber 5 and in the piston-side chamber 6 are equalized, and the actuator A generates a thrust force corresponding to a difference in pressure receiving area between the rod-side chamber 5 and the piston-side chamber 6 in the piston 3 multiplied by the pressure. In contrast, for the contracting operation of the cylinder body Cy, communication between the rod-side chamber 5 and the piston-side chamber 6 is blocked and the piston-side chamber 6 is brought into communication with the tank 7. Consequently, the actuator A generates a thrust force corresponding to the pressure in the rod-side chamber 5 in the piston 3 multiplied by the pressure receiving area on the rod-side chamber 5 side.

In brief, the thrust force generated by the actuator A corresponds to half the cross-sectional area of the piston 3 multiplied by the pressure in the rod-side chamber 5. Therefore, for controlling the thrust force of the actuator A, the pressure of the rod-side chamber 5 may be controlled both during the extending operation and the contracting operation. In the actuator A of the present example, since the pressure receiving area on the rod-side chamber 5 side of the piston 3 is set to half the pressure receiving area on the piston-side chamber 6 side, control is simplified because the pressures in the rod-side chamber 5 are equalized during both expansion and contraction when the same thrust force is generated during both expansion and contraction. In addition, since the amount of hydraulic oil with respect to the amount of displacement are also equalized, responses during both extension and contraction area advantageously equalized. Even when the pressure receiving area on the rod-side chamber 5 side of the piston is not set to half the pressure receiving area on the piston-side chamber 6 side, the thrust forces of the actuator A during both extension and contraction can still be controlled by the pressure in the rod-side chamber 5.

Again, the left end of the rod 4 in FIG. 2 and the lid 13 that closes the right end of the cylinder 2 are provided with mounting portions, not illustrated, such that the actuator A can be interposed between the vehicle body B and the truck T of the railway vehicle.

The rod-side chamber 5 communicates with the piston-side chamber 6 via the first passage 8, and the first opening-and-closing valve 9 is provided in the middle of the first passage 8. The first passage 8 communicates with the rod-side chamber 5 and the piston-side chamber 6 outside the cylinder 2, but may be provided in the piston 3.

The first opening-and-closing valve 9 is an electromagnetic opening-and-closing valve and has a communicating position for opening the first passage 8 and communicating the rod-side chamber 5 with the piston-side chamber 6, and a blocking position for blocking the first passage 8 and shutting off the communication between the rod-side chamber 5 and the piston-side chamber 6. The first opening-and-closing valve 9 is configured to take the communicating position when electricity is supplied, and take the blocking position when electricity is not supplied.

Subsequently, the piston-side chamber 6 communicates with the tank 7 via the second passage 10, and the second opening-and-closing valve 11 is provided in the middle of the second passage 10. The second opening-and-closing valve 11 is an electromagnetic opening-and-closing valve and has a communicating position for opening the second passage 10 and communicating the piston-side chamber 6 with the tank 7 by opening the second passage 10, and a blocking position for shutting off the communication between the piston-side chamber 6 and the tank 7. The second opening-and-closing valve 11 is configured to take the communicating position when electricity is supplied and take the blocking position when electricity is not supplied.

The pump 12 is driven by the motor 15 rotating at a predetermined rotating speed by being controlled by the controller C, and is a pump discharging the hydraulic oil only in one direction. A discharge port of the pump 12 communicates with the rod-side chamber 5 via a supply passage 16, and a suction port of the pump 12 communicates with the tank 7, and thus the pump 12 sucks the hydraulic oil from the tank 7 and supplies the hydraulic oil to the rod-side chamber 5 when driven by the motor 15.

As described before, the pump 12 only discharges the hydraulic oil in one direction and does not have to perform an operation for switching the direction of rotation. Therefore, the pump 12 has no chance of change in amount of discharge at the time of switching the rotation, and thus a cost effective gear pump, for example, may be used. In addition, since the direction of rotation of the pump 12 is always the same direction, the motor 15, which is a drive source for driving the pump 12, is not required to have high response for switching the rotation, and thus a motor of a reasonable price may be used as the motor 15. Note that a check valve 17 configured to prevent a back-flow of the hydraulic oil from the rod-side chamber 5 to the pump 12 is provided in the middle of the supply passage 16.

In addition to the configuration described above, the liquid pressure circuit HC of the present example is provided with a discharge passage 21 configured to connect the rod-side chamber 5 and the tank 7, and a variable relief valve 22 configured to be capable of changing a valve-open pressure in the middle of the discharge passage 21.

The variable relief valve 22 in this example is a proportional electromagnetic relief valve and is capable of adjusting the valve-open pressure according to the amount of current to be supplied, that is, minimizes the valve-open pressure when the amount of current is maximized and maximizes the valve-open pressure when no current is supplied.

In this manner, providing the discharge passage 21 and the variable relief valve 22 enables adjustment of the pressure in the rod-side chamber 5 to the valve-open pressure of the variable relief valve 22 during the extending and contracting operation of the cylinder body Cy, and thus enables control of the thrust force of the actuator A by the amount of current to be supplied to the variable relief valve 22. Providing the discharge passage 21 and the variable relief valve 22 eliminates the need for sensors required for adjusting the thrust force of the actuator A as well as the need for advanced control of the motor 15 for adjusting the discharging flow rate of the pump 12. Therefore, the cost of the railway vehicle vibration damping device 1 is reduced, and a secure system is achieved in terms of hardware as well as software.

Note that when the first opening-and-closing valve 9 is opened to close the second opening-and-closing valve 11, or when the first opening-and-closing valve 9 is closed to open the second opening-and-closing valve 11, the actuator A can exert a damping force for one of expansion and contraction only against vibration input from the outside irrespective of the driving state of the pump 12. Therefore, for example, when the direction of exerting the damping force corresponds to the direction of excitation of the vehicle body B by the vibration of the truck T of the railway vehicle, the actuator A may be functioned as a one-side effective damper. Therefore, the actuator A may easily realize semi-active control based on Karnopp Skyhook control, and thus may function as a semi-active damper.

Control of the valve-open pressure is facilitated by using the proportional electromagnetic relief valve that proportionally changes the valve-open pressure by the amount of current to be applied to the variable relief valve 22. However, any variable relief valve that can adjust the valve-open pressure may be used without limiting to the proportional electromagnetic relief valve.

The variable relief valve 22 opens the discharge passage 21 when an excessively large input is present in the extending and contracting direction to the cylinder body Cy and consequently the pressure of the rod-side chamber 5 exceeds the valve-open pressure irrespective of the opening-and-closing state of the first opening-and-closing valve 9 and the second opening-and-closing valve 11. In this manner, the variable relief valve 22 discharges the pressure in the rod-side chamber 5 to the tank 7 when the pressure in the rod-side chamber 5 reaches or exceeds the valve-open pressure, and thus prevents the pressure in the cylinder 2 from becoming excessive to protect the entire system of the actuator A. Therefore, providing the discharge passage 21 and the variable relief valve 22 enables protection of the system.

In addition, the liquid pressure circuit HC in the actuator A of this example includes a rectifying passage 18 configured to allow only a flow of the hydraulic oil flowing from the piston-side chamber 6 to the rod-side chamber 5 and a suction passage 19 configured to allow only a flow of the hydraulic oil flowing from the tank 7 to the piston-side chamber 6. Therefore, in the actuator A of this example, when the cylinder body Cy extends or contracts with the first opening-and-closing valve 9 and the second opening-and-closing valve 11 closed, the hydraulic oil is pushed out from the cylinder 2. The variable relief valve 22 applies resistance against the flow of the hydraulic oil discharged from the cylinder 2, and thus the actuator A of this example functions as a uniflow damper with the first opening-and-closing valve 9 and the second opening-and-closing valve 11 closed.

More specifically, the rectifying passage 18 communicates the piston-side chamber 6 with the rod-side chamber 5, is provided with a check valve 18a in the middle, and is determined to be a one-way passage allowing only a flow of the hydraulic oil flowing from the piston-side chamber 6 to the rod-side chamber 5. In addition, the suction passage 19 communicates the tank 7 with the piston-side chamber 6, is provided with a check valve 19a in the middle, and is determined to be a one-way passage allowing only a flow of the hydraulic oil flowing from the tank 7 to the piston-side chamber 6. The rectifying passage 18 may be concentrated to the first passage 8 assuming that the blocking position of the first opening-and-closing valve 9 is a check valve, and the suction passage 19 may be concentrated to the second passage 10 assuming that the blocking position of the second opening-and-closing valve 11 is a check valve.

In the actuator A configured in this manner, even when the first opening-and-closing valve 9 and the second opening-and-closing valve 11 takes the blocking position, the rectifying passage 18, the suction passage 19, and the discharge passage 21 connect the rod-side chamber 5, the piston-side chamber 6 and the tank 7 in communication in sequence. In addition, the rectifying passage 18, the suction passage 19, and the discharge passage 21 are set to be the one-way passages. Therefore, when the cylinder body Cy extends or contracts by an external force, the hydraulic oil does not fail to be discharged from the cylinder 2 and is returned back to the tank 7 via the discharge passage 21, and hydraulic oil shortfalls of the cylinder 2 is supplied from the tank 7 to the cylinder 2 via the suction passage 19. Since the pressure in the cylinder 2 is adjusted to the valve-open pressure by the variable relief valve 22 being resistant against the flow of the hydraulic oil, the actuator A functions as a passive uniflow damper.

At the time of a fail such that supply of electricity to respective devices in the actuator A is disabled, each of the first opening-and-closing valve 9 and the second opening-and-closing valve 11 take the blocking position, and the variable relief valve 22 functions as a pressure control valve having a valve-open pressure fixed to the maximum. Therefore, at the time of such a fail, the actuator A functions automatically as a passive damper.

Subsequently, in order to make the actuator A exert a thrust force in a desired extending direction, the controller C basically rotates the motor 15 and places the first opening-and-closing valve 9 to the communicating position while supplying the hydraulic oil from the pump 12 to the cylinder 2 and places the second opening-and-closing valve 11 to the blocking position. In this operation, the rod-side chamber 5 and the piston-side chamber 6 are brought into a communicating state, and thus the hydraulic oil is supplied from the pump 12 to both, while the piston 3 is pushed leftward in FIG. 2, and the actuator A exerts a thrust force in the extending direction. When the pressure in the rod-side chamber 5 and the piston-side chamber 6 exceeds the valve-open pressure of the variable relief valve 22, the variable relief valve 22 opens and the hydraulic oil is discharged to the tank 7 via the discharge passage 21. Therefore, the pressure in the rod-side chamber 5 and the piston-side chamber 6 is controlled to the valve-open pressure of the variable relief valve 22 determined by the amount of current applied to the variable relief valve 22. The actuator A then exerts a thrust force in the extending direction of a value corresponding to the difference in pressure receiving area between the piston-side chamber 6 and the rod-side chamber 5 of the piston 3 multiplied by the pressures in the rod-side chamber 5 and the piston-side chamber 6 controlled by the variable relief valve 22.

In contrast, in order to make the actuator A exert a thrust force in a desired contracting direction, the controller C rotates the motor 15 and places the first opening-and-closing valve 9 to the blocking position while supplying the hydraulic oil from the pump 12 to the rod-side chamber 5 and places the second opening-and-closing valve 11 to the communicating position. In this operation, the piston-side chamber 6 and the tank 7 are brought into a communicating state, and the hydraulic oil is supplied from the pump 12 to rod-side chamber 5, while the piston 3 is pushed rightward in FIG. 2, and the actuator A exerts a thrust force in the contracting direction. When the amount of current of the variable relief valve 22 is adjusted in the same manner as described above, the actuator A then exerts a thrust force in the contracting direction of a value corresponding to the pressure receiving area of the piston 3 on the rod-side chamber 5 multiplied by a pressure in the rod-side chamber 5 controlled by the variable relief valve 22.

Here, when the actuator A is not extended and contracted by an external force, but is extended and contracted by itself, the upper limit of the pressure in the rod-side chamber 5 is limited to the discharge pressure of the pump 12 driven by the motor 15. In other words, when the actuator A is not extended and contracted by the external force, but is extended and contracted by itself, the upper limit of the pressure in the rod-side chamber 5 is limited to the maximum torque that the motor 15 can output.

The actuator A functions not only as an actuator, but also functions as a damper only by opening and closing the first opening-and-closing valve 9 and the second opening-and-closing valve 11 irrespective of the driving state of the motor 15. Since troublesome and abrupt switching operation between the first opening-and-closing valve 9 and the second opening-and-closing valve 11 is not performed when the actuator A is switched from the actuator to the damper, a system of high response and reliability is provided.

Since the actuator A of this example is set to the single rod configuration, a long stroke length may be provided compared with the both-rod actuator. Therefore, the total length of the actuator is reduced, and improved mountability on the railway vehicle is achieved.

The flow of the hydraulic oil in the actuator A generated by the supply of the hydraulic oil from the pump 12 and the extending and contracting operation is configured to pass through the rod-side chamber 5, the piston-side chamber 6 in this order and finally be flowed finally back to the tank 7. Therefore, even when gas enters the rod-side chamber 5 or the piston-side chamber 6, the gas is autonomously discharged to the tank 7 by the extending and contracting operation of the cylinder body Cy, and thus deterioration of response in generation of the thrust force may be prevented. Therefore, troublesome assembly in the oil or assembly in the vacuum environment are not imposed in manufacture of the actuator A, and advanced deaeration of the hydraulic oil is not necessary in manufacturing the actuator A. Therefore, productivity is improved and the manufacturing cost is reduced. In addition, even when gas enters the rod-side chamber 5 or the piston-side chamber 6, the gas is autonomously discharged to the tank 7 by the extending and contracting operation of the cylinder body Cy. This eliminates the need for frequent maintenance for recovering the performance, and thus alleviates labor and cost burden in maintenance.

Subsequently, the controller C includes an acceleration sensor 40 configured to sense the lateral acceleration α of the vehicle body B, a target control force operating part 41 configured to obtain a target control force Fref that the actuator A is to be output, and a drive unit 42 configured to drive the motor 15, the first opening-and-closing valve 9, the second opening-and-closing valve 11, and the variable relief valve 22 based on the target control force Fref as illustrated in FIG. 3.

The acceleration sensor 40 senses the lateral acceleration α as a positive value when the lateral acceleration α faces rightward in FIG. 1 and senses the lateral acceleration α as a negative value when the acceleration sensor faces leftward in FIG. 1.

The target control force operating part 41 includes a bandpass filter 411 configured to filter the lateral acceleration α, a controller 412 configured to obtain a target control force Fref* for reducing the vibration of the vehicle body B from the lateral acceleration α, a bandpass filter 413 configured to filter the target control force Fref* before correction, a correcting part 414 configured to correct the target control force Fref* obtained by the controller 412 and output the final target control force Fref*, and a limiter 415 as illustrated in FIG. 3.

The bandpass filter 411 filters the lateral acceleration α to remove a stationary acceleration generating by the railway vehicle when traveling in a curve zone, a drift component and noise from the lateral acceleration α and inputs a result into the controller 412. The frequency band that the bandpass filter 411 allows to pass through is set to a value on the order of, for example, 0.5 Hz to 3 Hz to allow sufficient extraction of the acceleration component from 0.7 Hz to 2 Hz, which is a resonance frequency band of the vehicle body B while removing the stationary acceleration. In this manner, the stationary acceleration generating when traveling a curve, which is contained in the lateral acceleration α, is removed by the bandpass filter 411 and is input to the controller 412. Therefore, only the vibration which deteriorates riding comfort may be reduced.

The controller 412 is a He controller, and is configured to operate the target control force Fref* for reducing the lateral direction of the vehicle body B from the component of the resonance frequency band of the lateral acceleration α extracted by the bandpass filter 411. The component of the resonance frequency band of the lateral acceleration α extracted by the bandpass filter 411 is a vibration acceleration in the resonance frequency band in the lateral direction of the vehicle body B. Therefore, the target control force Fref* before correction obtained by the controller 412 is optimal for reducing the lateral direction of the vehicle body B.

The bandpass filter 413 in this example extracts only a frequency band in a range higher than 0.3 Hz, which is the frequency band of the stationary acceleration and lower than 1 Hz, which is a resonance frequency band of the vehicle body B. In other words, the bandpass filter 413 is configured to extract a low frequency control force Flow, which is a component of the frequency band from 0.3 Hz to 1 Hz contained in the target control force Fref* before correction.

As illustrated in FIG. 4, the correcting part 414 includes a gain setting part 4141 configured to set a value of a correction gain K based on the low frequency control force Flow obtained by filtering the target control force Fref* by the bandpass filter 413 and a gain multiplying part 4142 configured to obtain a final target control force Fref by multiplying the target control force Fref* before correction by the correction gain K.

The gain setting part 4141 compares the low frequency control force Flow output by the bandpass filter 413 with a threshold value Ft. When the low frequency control force Flow is the threshold value Ft or larger, the gain setting part 4141 reduces the value of the correction gain K, while when the low frequency control force Flow is smaller than the threshold value Ft, the gain setting part 4141 increases the value of the correction gain K. The correction gain K increases or decreases based on the reference described above in a range from the upper limit value 1 to the lower limit value 0.01. Specifically, when the low frequency control force Flow is the threshold value Ft or higher, the gain setting part 4141 counts time during which this condition is satisfied and decreases the value of the correction gain K during the time during which the condition is satisfied. Particularly, as illustrated in FIG. 5, time T during which the value of the correction gain K reduces from the upper limit value 1 to the lower limit value 0.01 is determined in advance. In other words, assuming that an operation cycle of the target control force Fref is t seconds, an amount of reduction β with respect to the operating cycle t of the correction gain K is determined in advance. Therefore, when the condition that the low frequency control force Flow is equal to or larger than the threshold value Ft is satisfied, the gain setting part 4141 obtains the value of the correction gain K by operating K=Kpre−β×t, where Kpre is a value of the correction gain K of the previous time. In contrast, when the low frequency control force Flow is smaller than the threshold value Ft, the gain setting part 4141 counts time during which this condition is satisfied, and increases the value of the correction gain K during the time during which the condition is satisfied. The time required for increasing the value of the correction gain K from the lower limit value 0.01 to the upper limit value 1 is set to be the same as the time required for reducing. In other words, assuming that an operation cycle of the target control force Fref is t seconds, an amount of increase with respect to the operating cycle t of the correction gain K is set to be the same value as an amount of reduction β. Therefore, when the low frequency control force Flow smaller than the threshold value Ft is satisfied, the gain setting part 4141 obtains the value of the correction gain K by operating K=Kpre+β×t, where Kpre is a value of the correction gain K of the previous time. Note that the time T required for maximum lowering and rising of the correction gain K may be tuned to be optimum for the railway vehicle, and for example, is set to a rage of approximately ¼ to ½ the cycle of the centrifugal acceleration during the travel in the relaxation curve zone.

As illustrated by a solid line in FIG. 5, when the time T is elapsed from a situation where the value of the correction gain K is 1 until the condition that the low frequency control force Flow is the threshold value Ft or larger is satisfied continuously, the value of the correction gain K is reduced from 1 to 0.01 little by little proportionally to time. When the value of the correction gain K reaches 0.01, the value is not reduced any more even when the condition that the low frequency control force Flow is the threshold value Ft or larger is satisfied. In contrast, after the value of the correction gain K reaches 0.01, and when the time T has elapsed in a state in which the conditions that the low frequency control force Flow is smaller than the threshold value Ft is satisfied continuously, the value of the correction gain K is increased from 0.01 to 1 proportionally to time. When the value of the correction gain K reaches 1, the value is not reduced any more even when the condition that the low frequency control force Flow is smaller than the threshold value Ft is satisfied. For example, when the condition that the low frequency control force Flow is equal to or larger than the threshold value Ft is satisfied continuously by time T1 (0<T1<T) from the state in which the value of the correction gain K is 1 is continuously satisfied, the value of the correction gain K is 1−β×(T1/t) Subsequently, when the condition that the threshold value Ft is smaller than the low frequency control force Flow is satisfied continuously for the time T1, the value of the correction gain K rises little by little to 1 in association with time as illustrated by a broken line in FIG. 5.

Note that the value of the threshold value Ft is set to a value corresponding to half the upper limit value Flim of the target control force Fref limited by the limiter 415. In other words, the threshold value Ft is set to Flim/2. The upper limit value Flim is set to a value that does not apply a load excessively to the motor 15 even when the actuator A continuously exert the upper limit value Flim.

The gain multiplying part 4142 multiplies the target control force Fref* before correction by the correction gain K set by the gain setting part 4141 to obtain a final target control force Fref and outputs a result to the drive unit 42. Therefore, since the correction gain K is reduced when the low frequency control force Flow reaches a value equal to or larger than the threshold value Ft, an opportunity that the final target control force Fref* after the correction reaches the upper limit value Flim is reduced.

The drive unit 42 includes a driver circuit configured to drive the motor 15, the first opening-and-closing valve 9, the second opening-and-closing valve 11, and the variable relief valve 22. The drive unit 43 causes the actuator A to exert a thrust force according to the target control force F by controlling the amount of current to be supplied to the motor 15, the first opening-and-closing valve 9, the second opening-and-closing valve 11, and the variable relief valve 22 in the actuator A according to the control force.

Note that the controller C may be provided specifically with, for example, an A/D converter for introducing a signal output from an acceleration sensor 40, a memory device such as a Read Only Memory (ROM) for storing a program to be used for a process required for importing the lateral acceleration α and controlling the actuator A, an operation device such as a Central Processing Unit (CPU) configured to execute a process based on the program, and a memory device such as a Random Access Memory (RAM) for providing the CPU with a memory area as hardware resources. The configurations of the respective parts of the controller C is achieved by executing a program for performing the process by the CPU.

A process to be performed by the controller C will be described with reference to a flowchart illustrated in FIG. 6. First of all, the controller C imports the lateral acceleration α (Step F1). Subsequently, the controller C obtains the target control force Fref* from the lateral acceleration α (Step F2). Next, the controller C performs a bandpass filter processing on the target control force Fref* to obtain the low frequency control force Flow (Step F3). Furthermore, the controller C compares the low frequency control force Flow and the threshold value Ft and determines whether Flow≥Ft is satisfied or not (Step F4). The controller C then reduces the value of the correction gain K when Flow≥Ft is satisfied (Step F5). In contrast, the controller C increases the value of the correction gain K when Flow<Ft is satisfied (Step F6). The controller C then obtains the final target control force Fref by multiplying the target control force Fref* by the correction gain K (Step F7). Finally, the controller C causes the actuator A to exert a thrust force by driving the motor 15, the first opening-and-closing valve 9, the second opening-and-closing valve 11, and the variable relief valve 22 based on a control force F (Step F8).

As described above, the railway vehicle vibration damping device 1 includes the actuator A interposed between the vehicle body B and the truck T of the railway vehicle and capable of exerting a control force; and a controller C configured to obtain the target control force Fref for reducing the vibration of the vehicle body B based on the lateral acceleration α of the vehicle body B, and the controller C includes a bandpass filter 413 configured to extract the low frequency control force Flow, which is a frequency component lower than the resonant frequency of the vehicle body B, and a correcting part 414 configured to correct the target control force Fref* based on the low frequency control force Flow.

As described above, when the railway vehicle travels from a straight zone to a relaxation curve zone provided between the straight zone and the stationary curve zone, a centrifugal acceleration having a frequency band higher than the frequency band of the stationary acceleration acts on the vehicle body B. In the bandpass filter 411 configured to filter the lateral acceleration α, the centrifugal acceleration during the relaxation curve zone cannot be removed. However, the railway vehicle vibration damping device 1 can extract a force component for resisting the centrifugal acceleration acting on the vehicle body B while traveling the relaxation curve zone from the target control force Fref* by the bandpass filter 413. The target control force Fref* is then corrected based on the low frequency control force Flow extracted by the bandpass filter 413, and thus the final target control force Fref may be prevented from becoming an excess value by an influence of the centrifugal acceleration during the travel in the relaxation curve zone, which cannot be removed by the bandpass filter 411.

In addition, although the target control force Fref* is corrected by the correcting part 414 based on the low frequency control force Flow, the target control force Fref* includes a component for reducing the vibration of the vehicle body B transmitted from the truck T side. Therefore, by limiting the target control force Fref after correction to be equal to or lower than the upper limit value Flim by the limiter 415, a force exerted by the actuator A is reduced, but a force component that reduces the vibration of the vehicle body B is not eliminated. Consequently, riding comfort is improved. Since the target control force Fref may be prevented from becoming an excess value by correcting the target control force Fref*, a large sized motor does not have to be used.

Therefore, according to the railway vehicle vibration damping device 1 of the present invention, a large sized motor is not required, and thus the cost and the mountability on the railway vehicle are not impaired. Therefore, the riding comfort during traveling in the relaxation curve zone may be improved.

Note that in this example, the bandpass filter 413 extracts a frequency component higher than the frequency band of the stationary acceleration of the target control force Fref* and lower than the resonant frequency of the vehicle body B. However, usage of the low pass filter that allows passage of the frequency component lower than the resonant frequency of the vehicle body B is also applicable. Even when the low pass filter is used in this manner, the target control force Fref* is obtained from the lateral acceleration α after removal of the stationary acceleration by the bandpass filter 411. Therefore, since the target control force Fref* is not increased by the influence of the stationary acceleration, downward correction of the target control force Fref* does not occur during traveling in the curve zone, and thus there is no disadvantage.

In addition, in the railway vehicle vibration damping device 1 of this example, the target control force Fref is corrected downward when the low frequency control force Flow is equal to or larger than the threshold value Ft by the correcting part 414. Therefore, an opportunity that the target control force Fref after the correction is limited by the limiter 415 is reduced.

In addition, in the railway vehicle vibration damping device 1 in this example, the correcting part 414 corrects the target control force Fref* by multiplying the target control force Fref* by the correction gain K having an upper limit value not larger than 1 when the low frequency control force Flow reaches the threshold value Ft or larger. In the railway vehicle vibration damping device 1 configured in this manner, since the frequency of the target control force Fref* before the correction and the target control force Fref after the correction has no change, an effect of reducing the vibration of the vehicle body B transmitted from the truck T is further enhanced. Therefore, according to the railway vehicle vibration damping device 1, riding comfort during the travel in the relaxation curve zone is further improved.

In the railway vehicle vibration damping device 1 in this example, the correcting part 414 reduces the correction gain K little by little while the low frequency control force Flow has the threshold value Ft or larger and increases the correction gain K when the low frequency control force Flow is smaller than the threshold value Ft. In other words, in the railway vehicle vibration damping device 1 in the example, the correction gain K is reduced according to time during which the target control force Fref* is equal to or larger than the threshold value Ft and is increased according to the time during which the target control force Fref* is smaller than the threshold value Ft. In this manner, according to the railway vehicle vibration damping device 1, the abrupt change in value of the correction gain K is alleviated, and thus the abrupt change of the target control force Fref after the correction is also alleviated.

In the description given thus far, the correcting part 414 reduces the correction gain K according to the time during which the target control force Fref* reaches the threshold value Ft or higher. However, the value of the correction gain K may be changed according to the amount of the target control force Fref* exceeding the upper limit value Flim. In other words, a configuration may be modified such that the larger the amount of excess of the target control force Fref* with respect to the upper limit value Flim, the smaller the value of the correction gain K becomes. Also, in the description given above, the correcting part 414 corrects the target control force Fref* by multiplying by the correction gain K. However, correction may be such that a constant value is subtracted from the target control force Fref* when the target control force Fref* is a positive value, and a constant value is added to the target control force Fref* when the target control force Fref* is a negative value. In this case as well, the frequency of the target control force Fref* does not occur before and after the correction.

Although the preferred embodiment of the present invention has been described thus far, alterations, modifications and changes may be made unless otherwise departing from claims.

The present application claims priority based on Patent Application No. 2016-176300 filed to Japan Patent Office on Sep. 9, 2016, and the entire part of which is incorporated by reference to the present description.

Claims

1. A railway vehicle vibration damping device comprising:

an actuator interposed between a vehicle body and a truck of a railway vehicle and capable of exerting a control force; and

a controller configured to control the actuator by obtaining a target control force for reducing a vibration of the vehicle body based on a lateral acceleration of the vehicle body,

wherein the controller includes

a filter configured to extract a low frequency control force, the low frequency control force being a frequency component lower than a resonant frequency of the vehicle body in the target control force, and

a correcting part configured to correct the target control force based on the low frequency control force.

2. The railway vehicle vibration damping device according to claim 1,

wherein the correcting part corrects the target control force downward when the low frequency control force is equal to or larger than a threshold value.

3. The railway vehicle vibration damping device according to claim 1,

wherein the correcting part corrects the target control force by multiplying the target control force by a correction gain having an upper limit value not larger than 1 when the low frequency control force is equal to or larger than a threshold value.

4. The railway vehicle vibration damping device according to claim 3,

wherein the correcting part decreases the correction gain little-by-little while the low frequency control force is equal to or larger than the threshold value and increases the correction gain when the low frequency control force is smaller than the threshold value.